Intermittent motion apparatus



Aug. 12, 1969 w. s. TOUCHMAN INTERMITTENT MOTION APPARATUS 13 Sheets-Sheet. 2

Filed Aug. 11, 1967 FIG. 4

H15 AT TORNEYS Aug. 12, 1969 w.s.1-oucHMAN 3,460,343

INTERMITTENT MQTIQN APPARATUS Filed Aug. 11, 1967 13 Sheets-Sheet L FIG.60

INVE N T WILLIAM S. TOUCHMAN HIS ATTORNEYS Aug. 12, 1969 w. s. TOUCHMAN INTERMITTENT MOTION APPARATUS l3 Sheets-Sheet 5 Filed Aug. 11, 1967 FIG. 8d

FIG. 33

INVENTOR WILLIAM S. TOUCHMAN HIS ATTORNEYS Aug. 12, 1969 w. s. TOUCHMAN 3,460, 3

INTERMITTENT MOTION APPARATUS Filed Aug. 11, 1967 15 Sheets-Sheet 6 N FIG. n N

W Distance Along Shaft a: a Q o"-SECT|ON NUMBERS INVENTOF! WILLIAM S TOUCHMAN s|o- 0F SHAFT INTO SECTIONS BY TORSlONAL ANALYSIS Q a ms ATTORNEYS Aug. 12, 1969 w. s. TOUCHMAN 3,

INTERMITTENT MQTIQN APPARATUS Filed Aug. 11, 1967 13 Sheets-Sheet 7 3 mm mm INVENTOR WILLIAM $.TOUCHMAN I q/ g 2 HIS ATTORNEYS mnf/ /////M/ //7 /JF// g- 12, 1959 w. s. TOUCHMAN 3,460,343

INTERMITTENT MOTION APPARATUS ms \\\\\\\\w WIL M S. CHMAN Max (MM/M HIS ATTORNEYS Aug. 12, 1969 w. s. TOUCHMAN INTERMITTENT MOTION APPARATUS.

13 Sheets-Sheet 9 Filed AIIS- IL 1967 FIG. l6

HAROENING LINEAR H SOFTENING FIG. I9

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Aug. 12, 1969 w. s. TOUCHMAN INTERMITTENT MOTION APPARATUS 13 Sheets-Sheet Filed Aug. 11. 1967 F I PRESS. I47.00 LBS. PER

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F PRESS. I 73.50 LBS. PER 50. IN.

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m w a 4 SSION RA INVE NTOR WILLIAM S. TOUCHMAN mas 1, sec.

Z- 12, 1969 w. s. ToucHMAN .3 3

INTERMITTENT MOTION APPARATUS Filed Aug. 11, 1967 13 Sheets-Sheet 12 INVENTOR WILLIAM S. TOUCHMAN O do! w R1,.

HIS ATTORNEYS 8- 12, 19559 w. s. TOUCHMAN 3 INTERMITTENT MOT ION APPARATUS Filed Aug. 11. 1967 13 Sheets-Sheet 18 I330 13; FIG. 30

HIS ATTORNEYS United States Patent 3,460,343 INTERMI'I'I'ENT MOTION APPARATUS William S. Tonchman, Dayton, Ohio, assignor to The N ational Cash Register Company, Dayton, Ohio, a corporation of Maryland Continuation-impart of application Ser. No. 609,378, Jan. 16, 1967. This application Aug. 11, 1967, Ser. No.

Int. Cl. F0311 1/00; Fl6f /16 US. Cl. 60-6 14 Claims ABSTRACT OF THE DISCLOSURE An intermittent motion apparatus for use in ultra high speed indexing. The input member of the apparatus is rotated at a constant velocity while the output member thereof is oscillated at a resonant frequency relative to the input member. Magnetically operated and fluid operated oscillators act upon a resilient means so as to cause said output member to effectively dwell a desired number of times for each revolution of said input member. Said resilient means includes a solid shaft which is placed in torsion, magnetic springs, and solid elastomeric elements, like rubber, which are operated in compressions and shear modes.

CROSS REFERENCES TO RELATED APPLICATIONS This application is a continuation-in-part of United States patent application Ser. No. 609,378, filed on Jan. 16, 1967, now Patent No. 3,389,843, said application being filed by applicant and assigned to the same assignee as the present application.

BACKGROUND OF THE INVENTION This invention relates to intermittent motion devices, and, more particularly, it relates to such devices which are capable of operating at ultra high stepping rates. These devices operate substantially at the resonant frequency thereof and are especially useful in devices such as high speed printers and tape punches although not necessarily limited thereto.

One of the prior-art intermittent motion devices includes a clutch means which is operatively connected between a rotatable input and a rotatable output means. When the clutch means is actuated, the input motion is transmitted to the output means, and, correspondingly, when the clutch means is deactuated, the motion of the output means ceases. By controlling the frequency of clutch actuation and deactuation, dilferent stepping or dwell rates can be obtained. The dwell rates in such a device are generally limited by the inertia of its moving parts when operating at high frequencies.

SUMMARY OF THE INVENTION Stated generally, the device of this invention has an input means which is rotated at a constant velocity and a rotatable output means which is operatively connected to the input means via a resilient means. Also included in this invention is an oscillator means which acts upon said resilient means at substantially the resonant frequency of the device so as to enable said output means to effectively dwell a desired number of times for each revolution of said input means while said input means is rotated at said constant velocity. The resilient means includes fluid means such as a gas, magnetic springs, and elastic members such as torsion shafts. The oscillator means includes oscillators which are fluid operated and those of the magnetic variety. Means are provided to keep the device in resonance and to control the amplitude of its oscillation, so that timeand position-regulated etfective dwell periods are obtained.

Experience with intermittent motion devices based on the disclosure of the previously named copending application, using torsion shafts as the resilient means at frequencies down to about twenty cycles per second, has demonstrated the following basic advantages for slower speed systems:

(a) Smooth and quiet operation;

(b) Low power requirements;

(c) Indexing accuracy unaffected by wear;

(d) Freedom from lubrication and maintenance; and

(e) Ease of manufacture.

Most of these advantages listed above were obtained from applicants devices of said copending application when said devices were operated at high frequencies; however, difiiculty was experienced at some of the coupling areas therein in that these areas appeared to be focal points for high frequency energy dissipation. As an example, in one of said devices, two pairs of jaw tooth couplings were used to detachably connect opposed ends of a print drum to the resilient means of the device, said print drum being the output member thereof. Evan though the tooth couplings were fabricated with the greatest precision, the contacting points of the tooth couplings appear to be focal points for high frequency energy dissipation. As a result, the oscillation of the drum or ouput member was somewhat unpredictable and erratic. In addition, some high frequency sounds appeared to emanate from vibrations normal to the surface of said print drum indicating that a method and a means were required to confine all vibrations to the resilient means used in the device. To aid in the eliminaton of such high frequency sounds, the subject of interface damping was studied, and the results thereof were incorporated in the embodiments of the instant application.

Other advantages of this invention are apparent in intermittent motion devices which are operated under conditions which require that the steady intermittent motion be inerrupted asynchronously and in such a way that the cyclic rhythm is not detrimentally disturbed. One such device is a multi-channel, paper tape punch which operates at a maximum rate of 250 cycles or punched rows per second. If the dwell time, during which punching takes place, occurs for a quarter of an operating cycle, as in a continuous harmonic motion type of system, the dwell time for said device corresponds to one millisecond, which is adequate for the punching operation. Because the coded information which is fed into such a paper tape punch is asynchronous in nature, the intermittent device which feeds the tape to the punch must also feed row after row of said tape in an asynchronous manner.

Additional advantages relating to linear and non-linear classes of resilient means used in the intermittent motion devices of this invention will become apparent upon reading the following description. The objects of the instant invention are:

(a) To provide intermittent motion devices which include asynchronous indexing;

(b) To provide intermittent motion devices which utilize non-linear resilient means and are capable of operating in resonance over a wide range of operating frequencies;

(c) To provide devices of the above type which are inexpensive to construct and maintain;

(d) To provide an intermittent motion device having a cylindrical drum as the output member thereof, the device being subject to a minimum of high frequency sound energy dissipation;

(e) To provide an intermittent motion device which is modular in construction;

(f) To provide an intermittent motion device which uses a fluid operated oscillator;

(g) To provide an intermittent motion device which uses a gas as the resilient means; and

(h) To provide an intermittent motion device which uses an improved magnetic oscillator.

DESCRIPTION OF THE DRAWINGS A description of the drawings is as follows:

FIGS. 1a to If inclusive are various geometrical models and graphs illustrating the principles used in developing this invention;

FIGS. 2a to 20 inclusive are cross-sectional views of a magnetic oscillator showing its rotor passing through different areas of the associated stator;

FIG. 3 is a cross-sectional view taken along the line 33 of FIG. 20 and shows a portion of the rotor construction;

FIG. 4 is a cross-sectional view taken along the line 4-4 of FIG. 2b and shows more details of the rotor construction;

FIGS. 5a to 5 inclusive are various geometrical models and graphs illustrating the principles used in developing this invention;

FIGS. 6a to 6d inclusive are cross-sectional views of a fluid-operated oscillator with its rotor in various positions relative to the discharge parts in the plenum;

FIG. 7 is a cross-sectional view taken along the line 77 of FIG. 2c and shows the relationship of the fluidoperated oscillator to a resilient means which is actually a torsion shaft;

FIGS. 8a to 8d inclusive are various geometrical models illustrating the principles used in developing this invention;

FIG. 9 is a geometrical model of a three-mass, twospring system used to illustrate the principles of this invention;

FIGS. low and 101) show a geometrical model and chart respectively, illustrating the princples used in this invention;

FIG. 11 is a chart used to illustrate the principles of this invention;

FIG. 12 is an elevational view, partly in cross section, showing one embodiment of the invention, which uses torsion shafts for the resilient means, a magnetic oscillator for the oscillator means, and a cylindrical drum for the output member of the device;

FIG. 13 is a view similar to FIG. 12 but shows only that portion thereof which relates to the resilient means connected to the cylindrical drum, and which view shows another embodiment of the resilient means;

FIG. 14 is an elevational view, partly in cross section, of another embodiment of this invention which uses metal plates as the resilient means to effect a modular-type construction;

FIG. 15 is an end view, partly in cross-section, and is taken along the line 15-15 of FIG. 14 to show the wedge members which secure the metal plates to the input shaft;

FIGS. 16 and 17 show various graphs used in illustrating the principles of this invention;

FIG. 19 is a cross-sectional view of another embodiment of the invention which is taken along a plane which is perpendicular to the common rotating axis for the input and output members and in which metal plates (the resilient means of the embodiment) wrap and unwrap upon carnrning surfaces associated with the rotating input member;

FIG. 20 is a chart used in illustrating the principles of this invention;

FIG. 21 is a geometrical model used in explaining the principles used in designing those embodiments of this invention which utilize a compressible gas as the resilient means;

FIGS. 22, 23, and 24 are various charts used in explaining the principles of this invention relative to those embodiments using a compressible gas as the resilient means;

FIG. 25 is a cross-sectional view of a first embodiment of the invention which uses gas as the resilient means and is taken along a plane which is perpendicular to the common axis of rotation for the input and output members;

FIG. 26 is a cross-sectional view similar to FIG. 25 but showing another embodiment of the invention;

FIG. 27 is a geometrical model used to explain some of the principles of this invention;

FIGS. 28 to 280 inclusive are views similar to FIG. 25 but showing different embodiments of the invention using a compressible gas as the resilient means;

FIG. 28d is a view similar to the embodiments of FIGS. 28a to 28d inclusive but showing an air bearing incorporated therein; and

FIG. 286 is an enlarged view of the portion of the drawing enclosed in the circle of FIG. 28d.

FIG. 29 is a graph illustrating some of the principles of the invention.

FIG. 30 is an elevational view, partly in cross section, of an embodiment of this invention which uses an elastomeric spring element which is operated in a shear mode.

FIG. 31 is a cross-sectional view taken along the line 3131 of FIG. 30, showing more details of the embodiment.

FIG. 32 is a cross-sectional view similar to FIG. 3], but showing an embodiment which uses an elastomeric spring element which is operated in a compressive mode.

FIG. 33 is a cross-sectional view of a drum and shaft of an embodiment of this invention which uses a magnetic spring.

DESCRIPTION OF THE PREFERRED EMBODIMENTS As stated earlier, high frequency energy dissipation was studied, and the results thereof were incorporated in the embodiments of the instant application, as it is important to keep both material damping and system or interface damping at a low level.

Material damping has been defined as the phenomenon by which energy is dissipated in a vibrating mechanical system consisting of a volume of macrocontinuous matter. The term macrocontinuous includes the damping originating at grain boundaries in metals, plastic slip or flow, dislocation movements, magnetomechanical effects, and inhomogeneous strain in fibrous materials. System or interface damping involves energy dissipation at interconnecting or distinguishable parts of an apparatus. A good example of system damping has already been given in relation to the jaw tooth couplings used to detachably connect opposed ends of a print drum to the resilient means used in the device. Even though these couplings were very carefully constructed and lapped in for almost perfect axial alignment, the mating surfaces of the tooth couplings began to show adverse wear effects, similar to fretting corrosion, after extended operation of the device.

The complete function of damping in the embodiments of this invention will be more readily understood if a systematic presentation of some of the analytical techniques for studying damping are reviewed. Experience with intermittent motion devices (based on said copending application and the embodiments of the present invention) over most of the useful frequency operating ranges, has indicated that all analytical analysis should be directed toward the steady-state conditions. Specifically, it has been found that the combined damping etfects of most of the systems investigated (that is, systems which have a rotary oscillation superimposed on a steady rotation) quickly restore the system to a steady-state condition following a step change of one or more of the input control parameters.

An analysis of damping problems related to an intermittent motion device of the type using a shaft subjected to torsional oscillations might begin with the following differential equation of motion, which applies to a linear system:

Equation 1 The symbols referred to in the above equation refer to those used in conjunction with the geometrical models shown in FIGS. 1a and 1b. For an intermittent device of the above type to be precisely linear implies that the effective driven inertial, the viscous damping coefficient 0, the spring constant k, and the peak torque amplitude L of the external (exciting) force, all remain constant. Assume first that the system under study has zero damping and is oscillating with a constant amplitude iqfi which means that the exciter term L sin wt is also zero. (The term L corresponds to the term used in said copending application.) With the above assumption, differential Equation 1 reduces to:

Equation 2 Iiia+kq9=0 The general solution of Equation 2 is:

Equation 3 in which the term /k/I refers to the angular natural natural frequency of the torsionally oscillating system, and in which Equation 4 and Equation 5 In the above equations, it is assumed that the inertia or mass of the spring (resilient member 1,) is zero, or is combined with the fixed inertia of the oscillating member (I Exact equations were developed in said copending application for taking into account the inertia of the spring (resilient member) itself.

FIG. 1c is a graph showing a typical force vs. deflection curve for a theoretically undamped oscillating member in which deflection and torque curves are defined by the following relations respectively.

Equation 6 Equation 7 L L sin wt torque.

If viscous damping and a harmonic excitation are added to the geometrical model shown in FIGS. la and lb to maintain a constant amplitude of oscillation, Equation 1 applies, and, considering only the steady-state solution when the oscillation is defined by Equation 6, this latter equation may be rewritten in terms of frequency as follows:

Equation 8 The net torque exerted on the inertia of I (FIGS. 1a and 1b) by the torsional spring (the shaft) and its assumed viscous damper is:

Equations 8 and 9 define the relationship between L and 9b, and, when specific values are used in the equations,

6 the results therefor may be plotted to produce the ellipse shown in FIG. 1d.

The energy dissipated in one cycle of oscillation is equal to the area of the ellipse shown in FIG. 1d and is:

When structural materials such as those which might be used for the torsion shaft of FIGS. la and lb are actually used, the assumption made regarding viscous damping is not valid, and the problem becomes more complex. In fact, the relation of energy dissipation to stress level introduces a nonlinearity into the system, producinga mechanical hysteresis loop (FIG. 1e) of increased area over that shown in FIG. 1d. An approximate analytical treatment can be achieved if one lets:

E==specific damping energy, equal to area within stressstrain hysteresis loop of the material, in.-lb./in. /cycle, and

E =total damping energy dissipated by the mechanical spring per cycle of vibration, in.-lb./ cycle.

Before the total damping energy E can be determined, the specific damping energy E must be related to the stress level 1-. A moderate amount of accumulated data for typical engineering materials illustrates the fact that the damping-stress relationship for all materials cannot be expressed by one simple function. For a large number of structural materials in the low-intermediate stress region (up to 70% of the fatigue strength at 2x10 cycles), the following relationship is reasonably satisfactory (from Shock and Vibration Handbook, edited by Cyril M. Harris and Charles E. Grede, pages 36 and 37, volume 2, McGraw-Hill Book Company, 1961):

Equation 11 E=C l" in which 0 and B are empirical constants and 1- is the stress level (lb./in. In general, values of 5:2.0 to 3.0 are common in the low-intermediate stress region but may be much larger at high stress levels.

In order to analytically determine the total damping energy E it is necessary to somehow solve the volumestress function of the mechanical spring element. Broadly, one must solve the following expressions:

Equation 12 V0 E =L EdV Equation 13 7 "r l 7 [sun wl cos mlv g +wl F An analysis of Equation 13 shows that the minimum energy design would theoretically be one for a torsion shaft having zero length. Practically, the minimum energy design for a material with this characteristic (3:2.0)

would be one of minimum shaft length consistent with maximum allowable stress.

For #:30

Equation 14 By differentiating Equation 14 with respect to length I, and setting the result equal to zero, the optimum shaft length for minimum energy is given by Equation 15 The optimum shaft length for minimum stress was given in said copending application as Thus,

Equation 16 mtn. stress min. enersY Equation 16 compares Equation 15 with the equation for minimum stress and shows that, if a material exhibits the damping property ti=3.0, the lowest energy dissipation per cycle will be achieved by shortening the shaft and allowing slightly higher stresses therein than that which would be produced with the minimum stress design. Using numerical integration techniques, any value of a and may be selected and introduced in the following equation for the determination of the damping energy per cycle:

Equation 17 Even if one assumes that the basic equation (Equation 11) applies, the derivation of Equations 13 through 17 is also made on the basis that and [i are constants from the axis of the torsion shaft to its extreme outer surface;

Equation 18 Damping present in the intermittent motion devices of said copending application and the instant invention rnust equate to the power input, which is equal to the sum of the following:

Damping losses (111) (a) Material damping, (b) Exciter damping, (c) Bearing damping, (d) Air damping, and (e) Interface damping.

In general, interface damping at the freqeuncy of the fundamental mode vibration must be virtually eliminated. If it is initially present in an intermittent motion device, it will, in all probability, increase as the device continues in operation. It is important that the lumped sum of the remaining damping parameters listed above remain essentially constant, as greatly simplified means for maintaining a constant amplitude of oscillation of the shaft or the output member is thereby made possible.

Such a simplified means of operating an essentially linear device at or near resonance can be understood by referring to FIG. If. A plot of r.p.m. vs. torque of a system excited for rotary oscillation is represented by a rise curve portion A, a fall portion B, and a peak point C. As fully described in said prior application, the curve ACB is in reality a frequency curve. since f=D r.p.m./60, in which D is the number of effective dwells per revolution of the forward rotation of the intermittent motion device. In operation, the system of FIG. 1 was maintained at a constant level of excitation. Operation at the resonant peak C in one embodiment of the invention corresponded to a steady-state input torque of approximately 158 in.-oz., with a motor speed of 1,710 r.p.m., and with a torsional oscillation amplitude of i018 radian.

The other curves on FIG. 1 labeled 6t), 70, 80, 90, 100, and (representing voltages), are speed-torque curves of an induction motor used to drive the oscillating system. The points where the motor speed-torque curves intersect the rise curve A of the vibrating system represent conditions of power balance; that is, the power output of the motor exactly balances the power required to overcome damping losses, or to maintain the corresponding peak-to-peak angle of the torsional oscillation. If the damping in the system and the motor characteristics therein remain constant, the angle of oscillation of the system will remain constant without the use of any control means.

If the motor voltage is increased to a point where its corresponding speed-torque curve (FIG. 1/) no longer intersects the rise curve A of the steady-state damping torque, the system immediately goes out of resonance with a sharp increase in motor speed. Intersecting points of the motors speed-torque curves with the fall curve B cannot be achieved, because, to operate successfully with the fall curve B, would require an electric motor whose output torque somehow managed to increase while the voltage thereto (to decrease its speed) was reduced. The foregoing method of operation for resonant oscillating systems is obviously not limited to induction motors, but any motor (series, D.C.. etc.) which has similar speedtorque curves, either designed for the application or under positive voltage control, can be used.

A second form of damping present in intermittent motion devices is air damping; however, the effect of air damping is relatively unimportant when compared to the other forms of damping present, and it may be assumed to be constant and of a very low value, at least for oscillating structures which expose a minimum of vibrating surface normal to an air column.

On the other hand, bearing damping can be large enough to be quite objectionable and must be considered in all practical designs. For example, the use of standard ball hearings on a shaft oscillating at frequencies between approximately 500 and 3,000 cycles per second with an angle of oscillation of approximately .018 radian has proven to be unsatisfactory. In such case, an angle of oscillation was reached beyond which the internal friction or damping losses of the bearing became excessive and unpredictable. The faulty performance of ball bearings under the high frequency-large angle oscillation appears to be caused by high forces created between the balls and the ball separator due to the high angular ac- (c) Hydrodynamic oil (journal bearing), and (d) Porous bronze, graphite, etc.

The hydrodynamic air or gas bearing, although workable, was omitted from the list because of high cost, difficulty in starting, lower reliability, and higher damping than the hydrostatic air bearing. Examples of some of the bearing types listed above will be included in the mechanisms to be described presently.

Exciter damping, or the damping present in the device which maintains the oscillating system at or near resonance, and at a constant amplitude, is the final system energy loss to be discussed. The construction of the exciter, or magnetic oscillator, was fully described in said copending application. The latter name was used in said copending application to identify a means for providing excitation at the resonant frequency of a torsionally vibrating system using either an electromagnet or a permanent magnet as the primary source for force generation.

The intermittent motion devices of said copending application and the instant invention are based on the excitation of a spring-mass system in which devices it is extremely diflicult to separate the exciter damping losses from the material damping losses of the spring or resilient element; however, since the primary concern is with the mechanical Q of the overall system, keeping these two damping losses grouped together presents no problem. For example, if an intermittent motion device includes a torsion shaft for the resilient member, and an oscillator which is magnetically operated, and if the design of said device is also such that the bearing damping and air damping related to the torsional oscillation at the fundamental frequency of the device can be neglected, then the power input of the mechanism during continuous intermittent operation thereof, minus the power input required for steady rotation without oscillation, equals the power required to overcome the losses in the spring member (material damping in torsion shaft), in the exciter (magnetic hysteresis), and at any interconnecting joints exposed to the fundamental frequency (interface damping).

Therefore, if the total damping power is known for a given system which is oscillating at a constant angleiqb at or near resonance, the approximate Q of the system can be determined by the following relations:

Equation 19 Q=21rW/ AW in which W is the stored energy in the vibrating system, and AW is the energy dissipated per cycle, as previously described. If the effective inertia I (lb.-in.-sec. the frequency f, and the angle of oscillation 5 (radians) are known, the stored energy in the vibrating system is given by:

Equation 20 P) max From Equation 8, the formula for angular velocity may be written as follows:

Equation 21 3:213. cos 2m From Equation 21, a maximum occurs when 1:0, and

Equation 22 (tun a. Substituting Equation 22 in Equation 20 yields Equation 23 W:21r If (in.-lb.)

If P equals the combined damping power when the peak angle of oscillation is iqb (Hp), then the energy dissipated per cycle due to damping is 10 Equation 24 AW (itL-lb.) Substituting Equations 23 and 24 in Equation 19 gives:

Equation 25 It has been shown that material damping increases at greater than a linear rate with increases in stress level; similarly, it has become apparent that in an electromagnetic exciter, damping loss increases at greater than a linear rate with increase in impressed coil voltage. The material damping losses in the spring or resilient element of a given system are determined by the steadystate angle of oscillation and its frequency. Power to sustain the steady-state vibration is supplied by the exciter, and its power level is precisely related to the damping loss power. Therefore, the material damping of the spring or resilient element also determines the damping of the exciter, and this fact makes the selection of the material for the resilient element of prime importance in the design of a system so that the combined damping effects will not be too great or too low for practical operation of the system.

Performance of the magnetic excitation devices disclosed herein has been quite satisfactory; however, for high frequency operation, another system, to be described, has certain advantages. These advantages will be described in relation to the specific embodiments disclosed herein.

After having reviewed some of the analytical techniques used in studying the problem of damping, it seems appropriate to describe how the subject of damping applies to the specific embodiments disclosed herein; however, before so doing, a general discussion of the various elements used in the intermittent motion devices will first be given.

FIG. 12 shows one embodiment of this invention which can be used to illustrate the general relationship of the various elements included therein. The output member of the device includes a drum 33, which is supported at opposed ends thereof by shafts 34 and 35, which in turn are rotatably supported in bearings 45 and 63, respectively. The drum 33 is fixed to said shafts 34 and 35 to rotate therewith. One end 34a of the shaft 34 is rotated by a motor 50, which is driven at a constant speed. In this embodiment, the resilient means employed includes the shafts 34 and 35, which are subjected to a torsional strain when the oscillator means is energized. The oscillator means shown in FIG. 12 comprises two magneticallyoperated oscillators 38 and 39, which, when energized, act upon the shafts 34 and 35, respectively, to restrain their rotation, thereby causing said drum 33 to periodically dwell at substantially the resonant frequency of the device a predetermined number of times for each complete revolution of said shafts 34 and 35.

While the embodiment of FIG. 12 has been described only generally, it serves to illustrate the general relationship of the basic elements included in the intermittent motion devices of this invention. In the descriptions of the specific embodiments which are to follow, only those elements of the devices which are novel will be described in detail; for example, FIGS. 2a to 20 inclusive, and FIGS. 3 and 4, show a magnetic oscillator which may be used in the general arrangement of the embodiments shown in FIGS. 12 and 14.

FIGS. 2a to 2c inclusive are cross-sectional views of a magnetic oscillator showing its rotor passing through different areas of its associated stator. The rotor includes a cylindrical element 1, which rotates between spaced opposed stator poles. This cylindrical element 1 may be utilized in an embodiment similar to that shown in FIG. 14, in which the cylindrical element is actually an extension of a drum 72.

Referring again to FIGS. 2a and 20 inclusive, the rotor and the stator of the oscillator shown therein are magnetically coupled as follows. The cylindrical element 1, which has a rotating axis 2, rotates in close proximity to the stator designated generally as 4. The stator 4 has an inner stationary, cylindrical member 5, whose axis is coincident with the axis 2, and which member is pro videcl with a plurality of pole faces 5a, which are equally spaced around its perimeter. The stator 4 also includes an outer stationary, cyclindrical member 6, whose axis is also coincident with the axis 2, and which member 6 is provided with a plurality of poles 6a, which are equally spaced around its inner wall so as to be in spaced opposed relation with said poles 5a to permit said cylindrical element 1 to freely rotate therebetween. The cylindrical element 1 is secured to an oscillating wheel or drum 3 (FIGS. 3 and 4), which also has a rotating axis coincident with the axis 2. The drum 3 is comparable in function to the drum 72, previously mentioned in connection with FIG. 14. The cylindrical members 5 and 6 of the stator 4 cooperate as shown in FIG. 4 to encase the coil 7, which is part of the magnetic coupling previously mentioned, said members 5 and 6 being made of magnetically soft material, so that the magnetic field strength between opposing pole pairs (5a and 6a) may be varied by controlling the voltage to the coil 7.

In the magnetic path of the oscillator shown in FIGS. 2a to 2c inclusive, and FIGS. 3 and 4, the cylindrical element 1 is provided with a plurality of conductor loops la, which are equally spaced around its perimeter. The cylindrical element 1 also has void areas 1b (FIG. 3) adjacent to said loops 1a, or said areas may be filled with a non-conductive material, so as to increase the torsional stiffness of the element 1, as may be required by its design.

The principle of operation relating to the magnetic oscillator just described is that a force is generated when a closed conductor loop is moved through a stationary magnetic field of varying intensity. In FIG. 2a, the conductor loops 1a are in a uniform magnetic field of maximum strength, and the voltage generated by the motion of the cylindrical element 1 in each of the separate conductors is equal, so that no current flows in the conductor loops la. A similar situation exists when the conductor loops in are in a field of minimum strength, as illustrated in FIG. however, when the conductor loops 1a are passing out of a magnetic field of maximum strength, as shown in FIG. 2b, unequal voltages are generated in each side of the individual loops 1a. As a result, unequal voltages are generated on each side of the loops, and the net voltage difference thereacross causes a current to flow in each loop, resulting in the creation of a drag force which opposes the motion of the cylindrical element 1. Likewise, when the conductor loops 1a are passing into a magnetic field of maximum strength (not shown), current is produced in said loops, so as to produce a force which opposes the motion of the cylindrical element 1.

The magnetic oscillator (FIGS. 21: and 2c inclusive) just described provides a simple means for producing the excitation necessary for some embodiments of the intermittent motion devices of this invention. As an illustration, the number of pole pairs (5a and 5b) of the stator 4, and the number of conductor loops (1a) of the rotor may both be conveniently made to equal one half the number of effective dwells (D/2) desired per revolution of the forward rotation of the input member of the device with which the magnetic oscillator is used. This feature will be more easily understood in connection with the description of FIGS. 5a to 5f inclusive.

A careful study of the various embodiments of this invention has revealed that, when the dwell therein lasts for approximately one quarter-cycle, the mathematical formula expressing this relationship can be written as:

Equation 26 =1rsec. 45"/4D=l.l072/D:l.ll072/D radians The above formula can be explained in relation to the geometrical model shown in FIG. 5a, which represents a stationary oscillating system. The system includes an inertia wheel (or drum) I which is secured to one end of a shaft k, whose remaining end is fixed. When the wheel 1,, is put in motion, it oscillates through an angle i as shown. The parameter D, which is the effective number of dwells per revolution of the wheel I is introduced into the mathematics of the system even though the wheel I does not rotate continuously in one direction. When D is plotted against cycle time ft, the graph shown in FIG. 5a results.

When a wheel or drum rotates continuously in one direction (b as shown in the geometrical model in FIG. 5b) and at a constant velocity, the following results obtain. The angular displacement 0 may be predetermined to be 21rft/D. When 0;- is plotted against the cycle parameter (ft), the graph as shown in FIG. 5b results.

The curve shown in the graph of FIG. 50 represents he sum of the curves shown in the graphs of FIGS. 5a and 5b, and the equation for this curve corresponds to the equation developed in said copending application, which is as follows:

Equation 27 Substituting Equation 26 in Equation 27 and multiplying through by D gives the equation for the normalized plot shown in FIG. 50:

Equation 28 6D:..1rt+l.1l072 sin 21 The curve shown in the graph of FIG. 5d represents a normalized plot of angular velocity vs. ft for the wheel or drum 3 (shown in FIGS. 3 and 4), and FIG. 5e represents a normalized plot of said drums acceleration. FIG. 5f is a plot of the exciter drag torque (T) vs. cycle time (ft) for the magnetic oscillator (FIGS. 2, 3, and 4) and shows how the high torque portions of the curve correspond to decreasing velocity portions of the velocity curve (FIG. 5d), thereby allowing the induction type exciter to effectively impart energy into the resonant system to balance the steady-state damping losses, as previously described. The waveform of the torque curve of FIG. 5f can be altered by changing the distance I between opposed portions of the conductor loops la (FIG. 3), or by altering the design of the pole faces 5a and 6a. or both. For proper action and greatest elficiency, the conductor loops 1a and the pole faces 5a and 6a must be evenly spaced around the center line 2.

The advantages of the magnetic oscillator shown in FIGS. 2, 3, and 4 over the magnetic type described in said copending application are that it adds less inertia to the wheel or drum inertia 1 and is more efficient at the higher frequencies.

A second modification of the oscillator means used in this invention includes an oscillator which is operated by a fluid (gas or liquid) and is shown in FIGS. 6a to 6d inclusive and FIG. 7.

The general arrangement of the fluid-operated oscillator relative to the intermittent motion devices of this inven tion is shown principally in FIG. 7. Only one end of the input shaft 10 for the device is shown, as the other end (not shown) is connected to to a source of constant rotation, as previously explained. The shaft 10 is actually the resilient means of the device and is subject to torsional strain when the fluid-operated oscillator is energized.

The fluid-operated oscillator shown in FIGS. 6a to 6d inclusive and FIG. 7 includes a stator 8 and a rotor 9. The rotor 9 is generally cylindrical in shape, has an axis of rotation which is coincident with the rotational axis of the shaft 10 (which may be a drum), and is secured to the shaft 10 by a fastener 11 for rotation therewith. The outer periphery of the rotor 9 has a plurality of equallyspaced projections therearound, each said projection having the curved surfaces and 9f shown in FIGS. 6a through 6d. The stator 8 is generally tubular in shape and envclops the projections of the rotor 9.

The stator 8 is actually a plenum 8a or chamber for the fluid (liquid or gas) which is used to actuate the oscillator. The plenum 8a includes outer cylindrical member 8b (FIGS. 6 and 7) and an inner cylindrical wall 8c, which are suitably sealed and are concentrically mounted relative to the rotor 9. The outer wall 8b is provided with an inlet 8 which is connected to a source of fluid pressure (not shown) to thereby deliver either a gas or a liquid to the plenum 8a. The inner wall 8c is provided with a plurality of groups of discharge ports 8d, through which the fluid is discharged. Each group of the ports 8d is arranged on a line which is axially aligned relative to the axis of the rotor 9, and the lines are equally spaced around the inner wall 80. The rotor 9 is also provided with axially aligned slots 9a, which are located between adjacent projections thereon to permit the fluid which has been used in the device to pass therethrough.

The fluid oscillator shown in FIGS. 6 and 7 operates as follows. One end of the shaft or drum 10 is supplied with a source of constant-velocity, rotary motion, as previously explained, which causes the rotor 9 to rotate relative to the stator 8. When the oscillator is energized, fluid under pressure passes from the plenum 8a through the discharge ports 8d and strikes the surface 9b of the projections of the rotor 9, giving it a forward thrust (counter-clockwise rotation, as viewed in FIGS. 6a to 6d inclusive), and, after the rotor rotates forward a short distance, the fluid from the discharge ports 8d strikes the opposite surfaces 9 giving the rotor a reverse or torque pulse. After striking the surfaces 9 and 9b, the fluid passes through the slots 9a, after which the fluid may be collected and returned to the pressure source by conventional recirculating means (not shown).

The number of groups of discharge ports 8d in the stator 8 is equal to D, and the number of combined surfaces 9 and 9b on the rotor 9 is equal to D, D being equal to the number of effective dwells per revolution of the input member in the device. If the axis relating to the plot of (ft) of FIG. 5 is moved to the line 0-0, it will represent the torque pulsation of the fluid-operated oscillator shown in FIGS. 6 and 7. By controlling the fluid pressure to the plenum 8a, the excitation power supplied to the resonant system can be regulated.

Torsionally oscillating systems conveniently divide into the following classes: (a) two-mass, one-spring systems; (b) three-mass, two-spring systems; (c) four-mass, threespring systems; and so on. FIGS. la and 1b show a twomass, one-spring system in which one of the masses or inertias is taken to be equal to infinity, since, as shown, the spring member (that is, the shaft) is attached to a rigid support. However, the instant application and said copending application are primarily concerned with a free-body, torsionally-vibrating system, with only the nodal points thereof undergoing a steady rotation.

Therefore, in reality, one is always concerned with a two-mass (or two-inertia) or larger system, because a rotating member with infinite inertia is not a practical reality. In particular, a system with a relatively high shaft inertia (or a high wheel inertia) or with a high combined shaft and wheel inertia will introduce an amplitude of oscillation into a flywheel by the following relation:

Equation 29 in which the subscripts f and w refer to the flywheel and the wheel, respectively, and which are taken to include the net effect of the distributed mass of the shaft. FIG. 8a shows how a flywheel system, using a torsion shaft 12 as a spring or resilient member, a wheel 13, and a flywheel 14, actually operates under torsional oscillation. As is apparent in FIG. 8a, the inertia I, of the flywheel 14 is much greater than the inertia of the wheel 13, and, during operation of the device with the excitation power on (the oscillator means not being shown), a

nodal surface, such as the surface of revolution of the approximate curve OAB, forms about the axis OX. This nodal surface is a surface of revolution only when the materials of the system are homogeneous and it is the only surface of the system which is in steady rotation or constant velocity rotation. In operation, the portion of the shaft 12 and the wheel 13 which are to the right of the nodal surface (formed by the curve OAB) oscillate opposed to the flywheel 14 lying to the left of the surface OABB', as viewed in FIG. 8a. Therefore, all points on the flywheel 14 to the left of said nodal surface will undergo a torsional, low-amplitude vibration. The length of the shaft 12 shown in FIG. 8a is obviously not the true acting length of the shaft 12; rather, a longer effective length 1 must be used to obtain the frequency of the system by calculation.

Also, as described in said copending application, the torsion shaft 1.2 shortens as the torsional shear stress therein increases. Since this decrease of the shaft occurs twice for each cycle of the torsional oscillation, a longitudinal vibration is set up which is two times the resonant frequency of the torsional oscillation. It is obvious that the longitudinal vibration of the wheel 13 is much greater than that of the flywheel 14, and, in cases where said wheels are solid masses, as shown in FIG. 8a, the amplitudes of their oscillations are approximately defined by Equation 29.

An interesting effect of the degree of amplitude of t rsional oscillation of the system is that a standard ball hearing may function properly at location 15 but quite unsatisfactory at location 16 of FIG. 8a.

Said copending application described a method of isolating the two inertias of the two masses (such as the wheels 13 and 14 of the model shown in FIG. 8a) from the effects of shaft shortening by the use of a diaphgram (as was done in the movie projector embodiment), but there are other spring systems which inherently do not produce this effect. For example, FIG. 8b, which is similar to FIG. 8a, shows a version of a compression spring used as a torsion member where the design of the spring can determine the rate of torsional stiffness to longitudinal stiffness. As illustrated, the spring would be fabricated from said bar stock; however, lower frequency systems could utilize conventional wire wound springs.

As the inertias of the wheel 13 and the flywheel 14 of FIG. 8a approach equality, the nodal surface OABB moves to the right and becomes a plane perpendicular to the axis of rotation when the inertias actually do become equal, as shown in FIG. 80. The nodal plane N-N lies midway between the wheels 17 and 18, which have equal inertias, and the torsional oscillations of said wheels are opposed relative to the plane N-N. Therefore, the system of FIG. 8c can be excited in two places namely, at the wheel 17 and at the wheel 18providing that there is a phase shift of degrees between the two exciters (not shown). As described in said copending application, the drive input for steady rotation can be applied at the nodal plane NN; however, if this is inconvenient, an alternate drive method can be used.

To understand this alternate drive method fully, it will be helpful to consider the torsional stiffness that is required for a system having a given frequency and number of dwells per revolution of forward rotation of the input member. FIG. 8d shows the arrangement of a system having two unequal inertias connected by one torsion spring 19, which has spring constants k and k which are effective from the nodal surface N--N (not a true plane) to wheel inertias 20 and 21, respectively. At far as the mathematics is concerned, the torsional oscillation on each side of the nodal surface NN is equivalent to the non-rotating system illustrated in FIG. 5a.

Referring to the geometrical models of FIGS. 1a and 1b, the following relationships apply:

l the wheel inertia with the effect of the shaft inertia included; and expressed in lb.-in.-sec. 

